FOLIA SCIENTIARUM
UNIVERSITATIS TECHNICAE RESOVIENSIS ZESZYTY NAUKOWE
POLITECHNIKI RZESZOWSKIEJ
NR 277
MECHANIKA
z. 81
(e-ISSN 2300-5211)
Wydano za zgodą Rektora R e d a k t o r n a c z e l n y Wydawnictw Politechniki Rzeszowskiej prof. dr hab. inŜ. Feliks STACHOWICZ
Zeszyt Naukowy Politechniki Rzeszowskiej Nr 277, Mechanika 81 został wydany przy współpracy:
Wydziału Budowy Maszyn i Lotnictwa Politechniki Rzeszowskiej Wydziału Matematyczno-Przyrodniczego Uniwersytetu Rzeszowskiego Wydziału Mechanicznego Uniwersytetu Technicznego w Koszycach, Słowacja
śytomierskiego Instytutu InŜynieryjno-Technicznego, Ukraina Opublikowane artykuły zostały zrecenzowane i zaakceptowane
przez Radę Naukową
K o m i t e t R e d a k c y j n y Władysław FILAR
Roman PETRUS Grzegorz PROKOPSKI
Jarosław SĘP Jan STANKIEWICZ
Zbigniew ŚWIDER
R e d a k t o r n a u k o w y Mieczysław KORZYŃSKI
p-ISSN 0209-2689 e-ISSN 2300-5211
Oficyna Wydawnicza Politechniki Rzeszowskiej ul. W. Pola 2, 35-959 Rzeszów
Nakład 180 egz. Ark. wyd. 11,53. Ark. druk. 9,75. Papier offset. kl. III 70g B1.
Oddano do druku w październiku 2010 r. Wydrukowano w październiku 2010 r.
Drukarnia Oficyny Wydawniczej, ul. W. Pola 2, 35-959 Rzeszów Zam. nr 101/10
RADA NAUKOWA I RECENZENCI ZESZYTU
Współprzewodniczący:
Krzysztof KUBIAK (Politechnika Rzeszowska) Józef TABOR (Uniwersytet Rzeszowski)
Miroslav BADIDA (Uniwersytet Techniczny w Koszycach, Słowacja)
Piotr P. MELNICHUK (śytomierski Instytut InŜynieryjno-Techniczny, Ukraina)
Członkowie:
Tadeusz BURAKOWSKI (Instytut Mechaniki Precyzyjnej w Warszawie) Jan BURCAN (Politechnika Łódzka)
Andrzej BYLICA (Uniwersytet Rzeszowski) Waldemar FURMANEK (Uniwersytet Rzeszowski)
Dusan KNIEVALD (Uniwersytet Techniczny w Koszycach, Słowacja)
Valerij KYRYLOVYCH (śytomierski Instytut InŜynieryjno-Techniczny, Ukraina) Juriy KOMAROV (Moskiewski Instytut Lotnictwa, Russia)
Mieczysław KORZYŃSKI (Uniwersytet Rzeszowski) Volodymir LIUBIMOV (Politechnika Rzeszowska)
Piotr P. MELNITSCHUK (śytomierski Instytut InŜynieryjno-Techniczny, Ukraina) Aleksander NAKONIECZNY (Instytut Mechaniki Precyzyjnej w Warszawie) Paweł PAWLUS (Politechnika Rzeszowska)
Piotr M. POVIDAYKO (śytomierski Instytut InŜynieryjno Techniczny, Ukraina) Włodzimierz PRZYBYLSKI (Politechnika Gdańska)
Stanisław PYTKO (Akademia Górniczo-Hutnicza w Krakowie) Jaroslaw SĘP (Politechnika Rzeszowska)
Laurentiu SLATINEANU (Uniwersytet Techniczny w Iasi, Rumunia) Emil SPIŠÁK (Uniwersytet Techniczny w Koszycach, Słowacja) Wiktor SZABAJKOWICZ (Łucki Uniwersytet Techniczny, Ukraina) Frantisek TREBUNA (Uniwersytet Techniczny w Koszycach, Słowacja) Krzysztof TUBIELEWICZ (Politechnika Częstochowska)
ZESZYTY NAUKOWE POLITECHNIKI RZESZOWSKIEJ Nr 277 Mechanika z. 81 2010
SPIS TREŚCI
1. Krzaczek P., Piekarski W.: Utilization of vehicle control-diagnostic system in eva-
luation of energetic parameters ... ...7 2. Kyrylovych V., Sazonov A.: Unit of adaptation grippers of industrial robots ... ...15 3. Lubas J.: A comparison of the tribological behaviours materials modified of boron
in the sliding pairs ………..………...…………. ...19 4. Malega P.: Model that rates technical effectiveness of production ... 1. ...29 5. Malejčík J., Brezinová J., Guzanová A.: High solid coatings quality in selected cor-
rosion media ... ...35 6. Malyarenko A., Mitenkov M., Kvasuk S.: Research of lap thermal deformation at
operational development of optical surfaces at the expense of a temperature varia-
tion of polishing slurry ... ...39 7. Olexová M., Slota J., Herditzky A.: Microstructure evaluation of trip steel used in
a car crash zones ... ...45 8. Pasternak I., Sulym H.: The unified approach for the analysis of elastic equilibrium
of solids containing thin internal and surface heterogeneities ... ...51 9. Shabaycovitch V.: Competitiveness of manufacture ... 2. ...61 10. Slota J., Gajdoš I.: Numerical simulation of the hydraulic bulge test of HSLA steel
and experimental verification of results ... ...67 11. Sobotová L., Spišák E.: Thermal drilling as a progressive technology of creating of
bushings ... ...73 12. Sobotová L., Dulebová L.: Evaluation of some mechanical properties of steel sheet ...79 13. Spišák E., Majerníková J.: Annealing process and its influence on mechanical
properties of packaging sheets ... ...85 14. Spišáková E.: The comparison of innovation activity of Slovak and Polish enter-
prises ... ...93 15. Todorov M., Draganov I.: Equations of elasticity theory in a helical coordinate
system ... ....103 16. Varga J., Greškovič F.: The influence of radiation crosslinking on mechanical
properties of plastics ... ....115 17. Viňáš J., Brezinová J., Guzanová A.: Tribological properties of selected ceramic
coatings ... ....121
6
18. Viňáš J., Brezinová J., Guzanová A., Lorincová D.: Hard surfacing repairing layers
in erosive wear process ... ....127 19. Viňáš J., Kaščák L., Ábel M., Draganovská D.: The quality analyze of MIG solde-
ring zinc-coated steel sheets by destructive testing ... ....135 20. Wolańczyk F.: The investigation of thermal conductivity of low-alloyed high speed
steels ... ....141 21. Zając G., Krzaczek P.: Comparison of rapeseed ethyl and methyl esters utilization
influence on engine energetic parameters ... ....145 22. Zdravecká E., Franta L., Ondáč M.: The carbon layers for bio-tribological aplica-
tions ... ....151 Index of authors ... ....155
Paweł KRZACZEK Wiesław PIEKARSKI
Uniwersytet Przyrodniczy w Lublinie
UTILIZATION OF VEHICLE CONTROL-DIAGNOSTIC
SYSTEM IN EVALUATION OF ENERGETIC PARAMETERS
In modern tractors and farm machinery newest electronic and information technologies, controlling not only engine work parameters but also operation of other sub-assemblies of the vehicle, are used. It enables real time exchange of information between controllers of tractor engine or whole agricultural unit. Evaluation of technical condition is possible on level of real time monitoring of diagnostic parameters, as well as by means of portable diagnostic systems. Utilization of on-board control-diagnostic systems, with installed automatic tests algorithms controlling operation of particular sub-assemblies, is also possible. Goal of this paper was evaluation of energetic parameters of John Deere 6920 tractor by means of on-board diagnostic system utilization. Tractor engine energetic parameters obtained as a result of such investigations enable evaluation of energetic
"saturation" of a tractor. Establishing characteristics of changes course for: torque, power and unitary fuel consumption enables determination of rotation speed range for optimal operation of an engine. Using on board diagnostics of agricultural tractor as well as additional external diagnostic system Service ADVISOR eases location of damages and failure repair in tractors and farming machinery. It also enables determination of unitary fuel consumption ge, precisely enough to be used in establishing its value for practical purposes.
INTRODUCTION
Analysing present situation and tendencies taking place in the word of motorization, increase in electronics, information techniques and mechatronics utilization becomes clearly noticeable It has significant effect on development of con- structional solutions as well as quality of operation of cars and commercial or agricultural vehicles. On one hand, it enables increase of work efficiency of exploited equipment, and on the other hand it leads to significant complication of construction [9]. Increasing number of various steering systems for particular subassemblies is introduced in vehicles It should be emphasised that areas of their operation often overlap, what results in necessity of integration of these system operation [1]. That is why, since 1980s, in cars, introduction of, based on various communication protocols, serial data buses has begun. Subsequently, unification of cars steering systems into common on-board diagnostic system (OBD), later adopted for commercial and agricultural vehicles, took place [2, 5, 10].
ZESZYTY NAUKOWE POLITECHNIKI RZESZOWSKIEJ NR 274 Mechanika z. 81 2010
8
Introduction of serial data bus, for example CAN type, enabled operation, on the same set of data, of even dozen or so controllers responsible for work of vehicle subassemblies. Utilization of data transmission protocols enables real time exchange of information between controllers. Moreover, its introduction simplifies assessment of cars technical condition [7,8]. It is especially important when exploited vehicles are scattered over significant area, and there are difficulties to use some of other measurement methods.
Utilization of serial data bus enables evaluation of technical condition of vehicle through monitoring diagnostic parameters in real time. Moreover, serial bus makes possible cooperation with external diagnostic system and is a basis for on-board control-diagnostic system, which is programmed to perform auto-tests of particular subassemblies [4].
Knowledge of energetic parameters, from the vehicle user point of view, is necessary to obtain good economics of vehicle work and is a basis for determination of its technical condition [6]. There are many methods of measuring energetic parameters such as power or torque. Some of them require transporting vehicle to particular place or disassembly of some of vehicle subassemblies. That is why, in this paper, method of measuring energetic parameters in place where vehicle is used was utilized. In order to measure fuel consumption, control-diagnostic system cooperating with mobile diagnostic system was used. Equipped with on-board diagnostic system, John Deere 6920 agricultural tractor was object of this research.
METHODS
In order to evaluate energetic parameters such as torque and power output of agricultural tractor, PT 301 MES device was used. It enabled measurements of these parameters on power take-off shaft (PTO) (Fig. 1). This device is a mobile dynamometric stand enabling measurements in vehicle workplace. The only limitation for its use is possibility of connecting it to a source of, depending on its load level, 400 or 240V current supply. Measuring range for maximal torque and maximal power, that can be absorbed by the brake, reaches 5800Nm and 340kW respectively, and depends on time of operation and conditions at the site.
Fig. 1. Diagram showing mobile PT 301 MES dynamometric station [4]
Investigated tractor reflects tendencies in modern construction solutions of vehicles. John Deere 6920 tractor is based on modular design. Its diesel engine is fully electronically controlled and equipped with eight, connected into common steering- diagnostic system, controllers managing particular subassemblies. The system enables, through diagnostic connector, two-way communication with external diagnostic system Service Advisor.
Before conducting research, equipment and technical condition of investigated tractor had been checked by means of visual inspection and interview with tractor user.
Next action was communication through diagnostic connector between engine steering units and portable computer equipped with diagnostic system Service ADVISOR.
After connection with on board diagnostic system was established, verification of previously gathered informations was conducted. Subsequently, it was determined what steering units are available in the system and what parameters can be monitored.
In order to take measurements, the tractor was placed on even surface and in line with the brake axis and power take-off shaft of the tractor was connected with a terminal of the brake by properly chosen transmission shaft. Following task was determining rotation speed ratio of the engine and the dynamometric stand by establishing ratio of rotational speed between the tractor engine crankshaft and power take-off shaft iwom = 1,995. Measurements of torque Mo and power Ne were carried out under full load and variable rotational speed. Readings of surrounding temperature Tot and atmospheric pressure pa were also registered.
During conducted research, in real time, following diagnostic parameters were registered in Service ADVISOR diagnostic system: hourly fuel consumption Gl [l·h-1], engine rotational speed , power take-off shaft rotational speed, air temperature Tot, atmospheric pressure pa, fuel temperature Tp, coolant temperature Tch and temperature of hydraulic oil Th. At the same time load, was put on the power take-off shaft by means of engine test bench PT 301 MES, and value of breaking power and rotational speed of the brake were registered.
Relying on above mentioned data, reduced torque Mozr and reduced power Nezr were calculated. Obtained maximal values of measured parameters were compared with producer specifications. Measurements of energetic parameters: power and torque were carried out in conformity with DIN 70020 standard, while calculating them to reduced conditions according ISO 3046 standard.
Measurements of hourly fuel consumption Gl were conducted simultaneously with measurement of torque Mo and power Ne. Obtained results of hourly fuel consumption Gl [l·h-1] were registered in Service ADVISOR diagnostic system. At the same time readings of air temperature Tot, atmospheric pressure pa and, in order to determine fuel density ρp, fuel temperature Tp were registered. Based on these measurements hourly fuel consumption Gc expressed in [kg · h-1] was calculated according to formula:
p l
c G
G = ⋅ρ [kg · h-1]; (1)
Than specific fuel consumption was determined according to formula:
n M
G N
g G
o c
e c
e ⋅ ⋅
= ⋅
= ⋅
π 2 1000
1000 [g · kWh-1]; (2)
10
where: Gc – hourly fuel consumption [kg · h-1], Ne – power output [kW], Mo – torque [Nm], n – engine rotational speed [rpm].
Moreover, flexibility coefficient e (3), enabling evaluation of suitability of vehicle for traction tasks, was determined. This coefficient is product of torque flexibility coefficient em (3) and rotational speed flexibility coefficient en (3).
M N
N n
m n
n M e M e
e= ⋅ = max ⋅ ; (3)
where: Mmax – maximal torque [Nm], MN – torque at maximal power [Nm], nN – maximal power rotational speed [obr · min-1], nM – maximal torque rotational speed [obr · min-1].
EXPERIMENTAL RESULTS AND DISCUSSION
Course of, registered in Service ADVISOR diagnostic system, hourly fuel consumption Gl and engine rotational speed n during taking measurements for John Deere 6920 tractor were presented in Figure 2. In order to make tractor measurement results analysis more accurate, measurement intervals were plotted on curves of course of both measured parameters, and whole measurement was divided into four periods.
Analogous analysis was used for evaluation of John Deere 6820 tractor [4]. Obtained results are presented in Table 1 and Figure 3.
Fig. 2. Progress of changes in hour fuel consumption Gl and engine speed n while preparing external speed characteristics for John Deere 6920 tractor (registered in the Service Advisor diagnostic system
In order to ensure proper measurements conditions, thermal state of engine, by means of registering temperature of coolant Tch, was monitored. Simultaneously, during measuring fuel consumption, temperature of fuel Tp, which was later used to recalculate volumetric fuel consumption (l·h-1) to, expressed in (kg·h-1), mass fuel consumption. Obtained results were presented in table 1.
First period included beginning of registration of parameters in Service ADVISOR diagnostic system and preparation of the investigated tractor for measurements. During this period, check-up of diagnostic parameters determining thermal condition of the engine and drive transmission system were conducted. In case of investigated vehicle these were: coolant, engine oil and hydraulic oil temperature. Second period, although the engine did not work with maximal possible fuel charge, represents increase of engine load with simultaneous increase of fuel consumption. Compared with investigation of John Deere 6820 tractor [4], this period comprised greater amount of measurement intervals, and rotational speed ranged from 2150 to 2250 rpm.
Table 1. Fuel consumption measuring results for John Deere 6920 tractor at maximum load and variable engine speed on the basis of data from the Service Advisor diagnostic system
Vehicle mileage P = 850 hours. PTO nominal rotational speedn = 1000 obr/min.
PTO ratio iWOM = 1,995. Weather conditions: pa = 1016 hPa, Tp = -4 °C.
Measured quantities Reduced quantities On. n
[rpm]
Mo
[Nm]
Ne
[kW]
Gl
[l·h-1] Tp
[ºC]
ρp
[kg·m-3] Gc
[kg·h-1] gp
[g·kWh-1] Moz
[Nm]
Nez
[kW]
Gc
[kg·h-1] gpz
[g·kWh-1] 1 2246 12,7 3,0 17,9 21,0 831 14,87 4986,8 12,1 2,8 14,87 5221,8 2 2210 253,6 58,7 20,5 21,0 831 17,04 290,2 242,2 56,1 17,04 303,8 3 2195 328,8 75,6 23,3 21,0 831 19,36 256,2 314,0 72,2 19,36 268,3 4 2167 407,6 92,5 27,3 21,0 831 22,69 245,3 389,3 88,3 22,69 256,9 5 2147 437,1 98,2 28,5 20,0 832 23,71 241,4 417,4 93,8 23,71 252,7 6 2123 447,0 99,4 28,4 20,0 832 23,63 237,8 426,9 94,9 23,63 249,0 7 2079 462,9 100,8 28,2 20,0 832 23,46 232,9 442,0 96,2 23,46 243,8 8 1983 481,4 100,0 28,0 20,0 832 23,30 233,0 459,8 95,5 23,30 244,0 9 1899 508,2 101,1 28,0 20,0 832 23,30 230,5 485,3 96,5 23,30 241,4 10 1803 534,4 100,9 27,7 20,0 832 23,05 228,3 510,4 96,4 23,05 239,1 11 1712 553,0 99,1 27,7 20,0 832 23,05 232,5 528,1 94,7 23,05 243,5 12 1624 557,1 94,7 27,1 20,0 832 22,55 238,0 532,0 90,5 22,55 249,2 13 1528 562,5 90,0 26,3 19,0 833 21,91 243,4 537,2 86,0 21,91 254,8 14 1444 565,2 85,5 25,8 19,0 833 21,49 251,4 539,8 81,6 21,49 263,2 15 1357 574,7 81,6 24,2 19,0 833 20,16 246,9 548,9 78,0 20,16 258,5 16 1261 572,5 75,6 21,7 19,0 833 18,08 239,1 546,7 72,2 18,08 250,4 17 1173 567,0 69,7 19,3 19,0 833 16,08 230,8 529,4 64,8 16,08 241,7 18 1081 534,4 60,5 17,4 19,0 833 14,49 239,5 358,1 41,0 14,49 250,8
Third period, similarly to one for 6820 tractor [4], included irregular course of both parameters. Increase of load governed by controller of dynamometric stand, causes variable reaction of the engine resulting from its work with maximal fuel charge, what makes it difficult to stabilize the load during the course of measurement. It should be noted that in case of 6920 tractor rotational speed range, for this period, was narrower (2050 - 2150 rpm), and variability of fuel consumption was significantly grater. It
12
enables statement that work of the engine with rotational speed grater than 2000 rpm is not beneficial. It should also be mentioned that significant increase of unitary fuel consumption occurred, after above mentioned rotational speed was exceeded.
Course of hourly fuel consumption Gl and rotational speed of the engine was stable in fourth period. In this period, rotational speed ranged from 1100 to 2000 rpm. For this interval, course of torque curve (Fig. 3) was flat, and at the same time specific fuel consumption was the lowest. However, analysing course of external characteristics of power Nez, torque Moz and specific fuel consumption, it was noted that the most favourable conditions for work of 6920 model are when rotational speed ranges from 1500 to 2000 rpm.
0 100 200 300 400 500 600
1000 1200 1400 1600 1800 2000 2200 n [rpm]
Moz [Nm], gez [g·kWh-1 ]
0 10 20 30 40 50 60 70 80 90 100
Nez [kW], Gl [l·h-1 ]
Moz gez Nez Gl
Fig. 3. Progress of changes in torque Moz, power Nez, hourly Gl and specific fuel consumption gez for John Deere 6920 tractor external speed characteristics on the basis of data from the Service Advisor diagnostic system
Obtained readings of minimal specific fuel consumption gemin for the investigated tractor were 239,3 g·kWh-1 for rotational speed of about 1800 rpm. Conducted research shows that minimal specific fuel consumption was obtained with the engine working with rotational speed for which maximal power is delivered. Comparison of obtained values of specific fuel consumption with research of Kamiński [3] (from 225 to 273 g·kWh-1) shows that for the investigated tractor specific fuel consumption was below mean values. Analysing course of specific fuel consumption curve it can be stated that optimal rotational speed, for this engine work, ranges form about 1500 to 2100 rpm. However, taking into consideration discussed before course of process of increasing load put on the power take-off shaft, this range should be narrowed to 2000 rpm.
In order to evaluate technical condition of the investigated tractor, comparison of obtained result with factory specifications for effective power Nwom at nominal range of revolution of power take-off shaft was conducted. In investigated vehicle, Nwom was 96,2 kW which is 97,5% of value stated by a producer. Therefore, vehicle technical condition can be considered good.
Results show that torque flexibility em = 1,13 and rotational speed flexibility en = 1,39, while flexibility of the engine as a whole e = 1,57. When compared with research of Kamiński [3], obtained coefficients are relatively low. It is related to course of torque and power curves, which in case of investigated vehicle have different character than in case of older generation vehicles, what results from a method of controlling work of ,powered with common-rail system, engine. It is clearly visible that nominal power is not equal to maximal power, and for rotational speed ranging from 1700 to 2100 rpm power is at similar level. Torque curve characterizes with flat course for rotational speed ranging from 1200 to 1700 rpm. Based on analysis of course of all curves it is possible to determine range of the engine work optimum rotational speed to be from 1500 to 2000 rpm.
CONCLUSIONS
Measurements of energetic parameters: effective power Ne, torque Mo, by means of utilization PT 301 MES dynamometric stand, in range of nominal rotational speed of power take-off shaft, enable precise evaluation of technical condition of investigated tractors in place where they are used. For the investigated vehicle difference was 2,5%, when compared with producer specifications , what proves its good technical condition.
Determining characteristics of torque, power and unitary fuel consumption course, and taking into consideration course of measurement registered in diagnostic system, enabled determination of engine work optimum rotational speed range, which, in case of John Deere 6920, was from 1500 to 2000 rpm.
Utilizing control-diagnostic system in vehicle with simultaneous use of external diagnostic system enables, accurate enough for practical purposes, determination of course of specific fuel consumption ge and its dependence on rotational speed. This system enables elimination of need for utilization of, interfering with engine powering system, research apparatus. That, thanks to lack of necessity of its installation, rationalizes measurement, and therefore, eliminates risk of accidental malfunctions. Its only necessary to carry out research which aim is to determine accuracy with which fuel consumption is measured by control-diagnostic system.
Moreover, introduction of diagnostic system, thanks to communication with vehicle on-board diagnostic system, contributes to significant facilitation of detection and removal of failures both in cars as well as tractors or agricultural equipment.
System enables identification of faults generated by tractor control systems, monitoring and recording of all vehicle work parameters. It can be used while servicing a vehicle, but it also enables, if necessary, analysis of recorded data in off- line mode.
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Inżynieria Rolnicza. Kraków 2007. Nr 6 (94). s. 57-63.
2. Jankowski M.: Wprowadzenie do pokładowego diagnozowania pojazdów samochodowych.
Artykuł XII Konferencji „Diagnostyka Maszyn Roboczych i Pojazdów”. Diagnostyka vol. 33.
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3. Kamiński J.R.: Analiza parametrów energetycznych ciągnika URSUS 1134. Inżynieria Rolnicza 3(91), 2007. s. 67-73.
4. Krzaczek P.: Assessment of energy parameters for John Deere 6820 farm tractor carried out using on-board diagnostics. Inżynieria Rolnicza. Kraków 2009. Nr 8 (117). s. 91-98.
5. Merkisz J., Mazurek S.: Pokładowe systemy diagnostyczne pojazdów samochodowych.
Wydawnictwa Komunikacji i Łączności. Warszawa 2007.
6. Rychlik A.: Metody pomiaru zużycia paliwa pojazdów użytkowych. Eksploatacja i Niezawodność.
Maintenance and Reliability. Nr 4/(32), 2006. s. 37-41.
7. Scarlett A. J.: Integred control of agricultural tractor and implements: a review of potential opportunities relating to cultivation and crop establishment machinery. Computers and electronics in agriculture, 30 (2001), 167-191.
8. Speckmann H., Jahns G.: Development and application of an agricultural BUS for data transfer.
Computers and Electronics in Agriculture, 23, 1999. str. 219-237.
9. Zhang Q.: Mechatronics and its applocation to off-road vehicels. Strona internetowa:
http://ageweb.age.uiuc.edu/faculty/research/mechatronics.htm. (2003)
10. Zimmermann W., Schmidgall R.: Magistrale danych w pojazdach. Protokoły i standardy.
Wydawnictwa Komunikacji i Łączności. Warszawa, 2008.
Valeriy KYRYLOVYCH Artyom SAZONOV
Zhytomyr State Technologycal University, Ukraine
UNIT OF ADAPTATION GRIPPERS OF INDUSTRIAL ROBOTS
In the paper presents a new approach to decision the problem of adaptation grippers of industrial robots to changing negative force-torque loads at the time of technical robot- ized kit fixing in the device of position working. Capabilities and basic elements of unit of adaptation grippers for industrial robots presented.
INTRODUCTION
One of the major problems with the use of industrial robots (IR) in machining flexible automated production is a problem of adaptation grippers IR (GrIR) to changing negative force-torque loads arising from the adjustment of each t-th position working PWt-1 at the time of fixing technical robotized kit (TRK), which includes long handling object HOt-1, which is clamping in grippers of IR (GrIR+HOt-1), clamping elements PWt accessories.
The interaction of technological items robotized system: on one side – the influence of a force HOt GrIR clamp on the other – force clamp HOt GrIRt-1 is called “dual connection”
[5]. Intensity clamp HOt-1 device of PWt far outstrips the amount of clamp force HOt by de- vice of PWt, which is obvious. The result of this interaction is force conflict (FC) [7].
Analysis of information sources showed that over the research of many researchers in- volved. Proposed some methods to adapt ShPR FC phenomena represented in many infor- mation sources [5, 6, 7, 8]. Methods of adaptation are mostly passive in nature.
THE AIM
The aim of this work is to present the proposed construction the unit of adaptation GrIR to FC, which is mounted on GrIR and reduces the negative impact of force-torque excitation on elements of IR.
The analysis of existing methods of adapting past can be divided into the following types [2, 3, 6, 8]: passive, active, combined and integrated.
Passive adaptation – involves in the construction GrIR compensating mechanical de- vices that react to the occurrence of unwanted effects force-torque and due to the ductility of structural elements compensate for installation errors that arise while fixing HOt-1 in device of PWt. This kind of adaptation of a number of disadvantages that reduce the possibility of ZESZYTY NAUKOWE POLITECHNIKI RZESZOWSKIEJ NR 274 Mechanika z. 81 2010
16
using it to systems that are moving manipulation of IR occurs at high speeds, due to pres- ence of unwanted vibration ultimate level of manipulation, the need to re-establish the com- pensation of errors when serving each PWt, but at the same time adapting this method is very simple in implementation.
Active adaptation – is realized by practicing corrective movements all degrees of mobil- ity IR. This kind of adaptation is limited primarily kinematic features junctions parts of IR manipulation, relatively large energy consumption of corrective movements drive train parts, requires certain hardware and software control systems improvements IR. Indisput- able advantage of this type is reliability adaptation and lack of critical loads on the elements and components of IR manipulation.
To combine the advantages outlined above approximation methods suggested to use the combined type of adaptation that does not require improvements IR and control systems through the use of mechanical devices compensating simplifies implementation and opera- tion.
THE PRESENTATION OF THE MAIN PART
As an example hardware implementation of the combined type of adaptation method presented unit of adaptation (UA), which conditionally divided into two blocks: Mechani- cal and Information. Mechanical block compensating role played by the compliance unit, which includes compensators angular and linear errors installing HO to avoid unwanted vibrational processes in the endpoint trajectory displacement of IR manipulation unit compliance mechanism for fixing the fixed position (MFFP). Pliability joints secured by using their spring design elements.
Information recording unit provides information about available installation error, converting mechanical quantities of forces and moments in the electric signal, further processing and signal control output to enable or disable the MFFP. The specified func- tion is achieved by using force-torque sensor as a sensitive and perceptive about forces and moments that occur during consolidation HOt-1 device of PWt clamping elements.
Sketches of the basic units and units represented in Figure 1.
Fig. 1. Basic elements of UA GrIR to the effect of FC
Moving TRK (GrOR + HOt-1) in the service area of device of PWt compliance re- corded MFFP unit that provides opportunities to move at high speeds and at the same time avoid unwanted vibrational processes in endpoint positioning GrIR. When clamp- ing HOt-1 device of PWt HOt-1 attached force-torque to load from the clamping device elements. Moment of load application is fixed information block, which produces sin- ghalese unfastening of MFFP to provide certain elements compensating compliance.
Thus, since the vast majority are device of PWt force-torque, is centering and setting compensation of errors. Change force-torques loads on HOt-1 shows the passage of time fixing process, where the change signal from the sensor is constant. This suggests that the transition is set to take place, all the error installing themselves compensated, so an information pack produces signal to turn on MFFP and recording site in compli- ance compensated position, the final link of IR removed from service area device of PWt.
CONCLUSIONS
Using the combined method of adapting the example of the proposed UA increases flexibility of automated production technology enables to increase the functionality of the conversion process in establishing long HO in device of PWt.
REFERENCES
1. Бауманн Э.: Измерение силы электрическими методами, пер. с нем. А. С. Вищенкова и С.
Н. Герасимова, под ред. И. И. Смыслова, М.: Мир 1978. 430 с.
2. Кирилович В.А., Черепанська І.Ю., Сазонов А.Ю.: Адаптивність схватів промислових роботів як напрям підвищення ефективності роботизованих механоскладальних технологій . Вісник ЖДТУ. Житомир, 2010 р. № I(52). С.17 – 24.
3. Кирилович В.А., Черепанська І.Ю., Сазонов А.Ю.: Адаптивність схватів промислових роботів механообробних ГВК. Методи розв’язування прикладних задач механіки деформованого твердого тіла. Збірник наукових праць Дніпропетровського національного університету. Дніпропетровськ, 2010 р. № 11. С. 119 – 125.
4. Проць Я.І.: Захоплювальні пристрої промислових роботів: [навчальний посібник].
Тернопіль, Тернопільськй державний технічний університет ім. І. Пулюя, 2008. 232с.
5. Тютюнник А.Г., Кирилович В.А., Чевпотенко О.В.: До питання адаптивності захватних пристроїв промислових роботів при синтезі роботизованих механоскладальних технологій.
Вісник ЖДТУ. Житомир, 2004 р. № IV(31). С.168 – 173.
6. Coiffet Ph.: Interaction with the environment. Robot technology. Volume 2. London, Kogan Page Ltd., 1983, – 240 с.
7. Monkman. G. J., Hesse. S., Steinmann. R., Schunk. H.: Robot grippers. Weinheim: WILEY-VCH Verlag GmbH & Co. KGaA, 2007. – 452 с.
8. Xiong Ch., Ding H., Xiong Y..: Fundamentals of robotic grasping and fixturing. USA: World Sci- entific Publishing Co. Pte. Ltd., 2007 – 229 c.
18
Janusz LUBAS
Rzeszow University, Poland
A COMPARISON OF THE TRIBOLOGICAL BEHAVIOURS MATERIALS MODIFIED OF BORON
IN THE SLIDING PAIRS
The aim of the present work is to determine the influence of technologically produced boron surface layers on the friction parameters in the sliding pairs under lubricated friction conditions. The tribological evaluation included ion nitriding, powder-pack boronizing, laser boronizing, hardening and tempering surface layers and TiB2 coating deposited on 38CrAlMo5-10, 46Cr2 and 30MnB4 steels. Modified surface layers of annular samples were matched under test conditions with counter-sample made from AlSn20 bearing alloy. Tested sliding pairs were lubricated with 15W/40 Lotos mineral engine oil. The tribological tests were conducted on a T-05 block on ring tester. The applied steel surface layer modification with boron allowed creating surface layers with pre-determined tribological characteristics required for the elements of sliding pairs operating under lubricated friction conditions. Boronizing reduces the friction coefficient during the start-up of the frictional pair and the maximum start-up resistance level is similar to the levels of pairs with ion nitride surface layers.
1. Introduction
The surface of engineering components are subjected to higher stresses and greater fatigue, abrasion and corrosive damages than the interior. Therefore, more than 90 pct of the service failures of engineering components are initiated at the surface. Surface modification techniques are employed to improve the resistance to failure by produ- cing a hard and wear resistance layer around a soft and tough core. Two major classes of treatments available for enhancing the surface properties are thermal and thermochemical. Thermal treatment, such as flame and induction hardening, modify the microstructure without modifying the surface chemistry, whereas in thermochemi- cal methods, the surface chemistry is altered. Nitriding and quenching and tempering are well know surface treatment methods [1].
The boride layer is formed by diffusion of boron atoms in the base metal at high temperature [2-5]. Diffusion of boron into the surface of various metals and alloys results in the formation of metallic borides which provide extremely hard (up to 2000HV), wear and corrosion resistant surface [6, 7]. This treatment is a thermoche- mical treatment in which a material is kept in a boron-giving environment from 850 to 1000oC for 2-10 h [3, 8]. Various processes adopted for boronizing include pack boriding, molten salt boriding, vacuum boriding, laser boriding, etc. The resulting layer may consist of either single-phase boride (FeB or FeB2) or polyphase boride layer (FeB and FeB2). The tribological properties variations of FeB and FeB2 layers depend on physical state of boride source used, boronizing temperature, treatment ZESZYTY NAUKOWE POLITECHNIKI RZESZOWSKIEJ NR 274 Mechanika z. 81 2010
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time, and properties of the boronized material [9]. Industrial boriding can be carried out on most ferrous and non-ferrous materials. Boronizing is very effective, especially on grey and ductile iron or low alloy steel with chromium. If boronizing is applied to a material surface the resultant boride layer increases the wear resistance considerably.
Furthermore, it decreases the friction coefficient [8]. The use of ceramic materials for many tribological applications has increased considerably over the past two decades.
This is the effect of the unique combination of properties of these materials, such as low bulk density, high corrosion resistance, low thermal expansion and high hardness over a wide range of temperature. Titanium diboride (TiB2) belongs to this group of materials and is well known for its high hardness, high melting point, a relatively high strength, high chemical stability at high temperature and high wear resistance [10, 11].
Titanium diboride is promising class of advanced materials which have a great potential for tribological application.
The current boronizing processes allow obtaining surface layers of high hardness and high resistance to corrosion and wear, with low brittleness and no tendencies towards cracking [4]. However, the operation characteristics of these layers depend on the chemical composition, the structure of the surface layer, the method and parameters of their production, as well as any possible thermal treatment. The modification of the surface layer with boron should be selected upon the required operating characteristics and the operating conditions of the kinematic sliding pairs [5- 7]. Thus, it is crucial to determine the influence the boron modification the elements of the sliding pair has on the operating conditions and wear during the lubricated friction.
2. Experimental
The aim of this work is to determine the influence of technologically produced boron surface layers on the friction parameters in the sliding pair. The tribological tests were conducted on a T-05 block on ring tester (Fig. 1). Three types of steel were used in the creation of annular samples, 38CrAlMo5-10, 46Cr2 and 30MnB4 (Fig. 1).
a) b)
Fig 1. Simplified scheme of the block-on-ring tester (a) and the sliding pair; 1-annual sample, 2- counter-sample (b)
Annular samples from 38CrAlMo5-10 steel were ion nitrided in the atmosphere H2 + N2, in the temperature of 500oC and in the time of 6h. Annular samples from 46Cr2 steel were borided in powder, in the temperature of 950 oC in the time of 8h, and then were isothermally quenched and hardened. In the boronizing process, powder of the following composition was used, B4C (30%), Al2O3 (68%), NH4Cl and NaF).
Annular samples from 46Cr2 steel were also laser-borided, with the use of CO2 laser (power of beam P = 2 kW, spot diameter d = 4 mm, energy density 160 W/mm2,
1 2
tracking speed v = 16 mm/s, gas carrier –argon). The boronizing process consisted in covering the annular sample with the layer of amorphous boron and liquid glass, and melting with the laser beam. Also, the annular samples from steel 46Cr2 were covered with the TiB2 coating using PVD method (temperature 400oC, time 40min, pressure in ionization chamber p = 2,5 x 10-2 bar). Annular samples from 30MnB4 steel were hardened and tempered, hardening in the temperature of 800oC and drawing temper in the temperature of 450oC. Modified surface layers of annular samples were matched under test conditions with counter samples made from AlSn20 bearing alloy. Tested sliding pairs were lubricated 15W/40 Lotos mineral engine oil.
The on-site tests were executed by following a specified algorithm, which included the initial grind-in of the samples and the correct co-operation process at the pre-determined load parameters. The grind-in was executed on a test site at a load of 5 MPa until a complete adhesion of the annular sample and the counter-sample was achieved. It was assumed for the starting phase that the pair would be accelerated from a speed of n = 0 to 500 rpm in 30 seconds. The friction coefficient to temperature ratio was measured as a function of pressure, while the measurements of the pair co- operation under the pre-determined friction conditions were taken at the annular sample rotational speed of 100 rpm and at a jump unit pressure p = 5, 10, 15 and 20 MPa. The course of the friction coefficient, temperature and wear as a function of variable load were registered at real-time during the tests.
3. Results and discussion
The co-operation of a sliding pair is characterised by the large dynamics of the measured parameters’ values, due to external forces. Determination of these changes’
tendencies is especially important in the start-up stage of the frictional work. The assessment of the occurring changes is possible by registering the friction coefficient as a function of variable sliding speed (Fig. 2). The registered charts present the typical courses of the friction coefficient for the ‘ring-block’ frictional pairs under the load of 20 MPa. During the first start-up phase, a rapid increase in the frictional resistance occurs, followed by its significant drop. The registered courses of the friction coefficient for the higher sliding speeds are diversified. There are sliding pairs, with an increase in the sliding speed of the annular sample, which cause the increase in the friction coefficient. These variations occur in the pairs with nitrided surface layers, after exceeding the sliding speed of 0.6 m/s and after 0.2 m/s for the pairs with the TiB2 coating. The measured value of the friction coefficient level in the associations with the laser borided layer and TiB2 coating equals approximately 0.11. As for the pairs with 30MnB4 steel annular samples, an increase in the sliding speeds leads to the stabilisation of the friction coefficient values, while the powder-park boronizing pairs exhibited a decrease in this value (to its lowest possible level µ = 0,02) along with an increase in the sliding speed.
Another significant aspect pertaining to sliding pairs is to determine the value of the start-up moment (Fig. 3). During the tests, the lowest friction resistances were recorded for the pairs with ion nitrided, powder-pack borided and hardened and tempered surface layers, which have similar values of approximately 8 Nm (at 20 MPa). Significant increases of the friction moment (by about 20%) occur in pairs with TiB2 coating and laser borided surface layer of annular samples. Similar moment
22
changes are observed at the pressure of 10 and 15 MPa. The pairs with nitrided surface layers and TiB2 coating reach the friction moment value at about 2 Nm under singular pressures (at 5 MPa), while the pairs with laser and powder-pack borided surface layers have the friction resistance values higher by 15%. The observation of the friction parameter changes during the start-up phase tells about the behaviour of the system during its further work. The most favourable operating conditions are present in sliding pairs in which the friction coefficient increases in the initial stage of start-up, and then decreases significantly and stabilises itself at a constant level. The value of the moment determines the energy demand of the system upon its start-up. Those sliding pairs which exhibited the tribochemical equilibrium within the shortest time generate optimal conditions for their further operation. The changes registered are the result of physiochemical processes and the changes in the friction surface micro- geometry due to the adaptation of the system to the conditions of external forces [12, 13]. In the sliding pairs, which exhibit a significant decrease in the friction coefficient, the improvement in the friction conditions depends on the increase in the effectiveness of the lubrication by the oil coat, due to the existing tribochemical changes. These changes are shaped by the existing load state of the kinematic sliding pair, the temperature levels and the chemical reaction occurring within the area of friction. As an effect of the changes in the oil chemical composition and the synthesis of new chemical compounds, a boundary layer is created, which strengthens the anti-wear layer by changing its structure and decreases the movement resistances. These changes lead to the further decrease in the friction resistance, accompanied by the increase in the sliding speed of the annular sample [12, 13]. The pairs with stable courses of the friction coefficient the surface layer of the element provides sliding characteristics, which allow the equilibrium of tribochemical phenomena within the contact area. This equilibrium allows inherent regulation of the processes occurring within the friction area, which stabilises the resistance values despite the increase in the sliding speed.
Fig. 2. Influence of surface treatment annular sample on change of friction coefficient vs.
rotation speed and load 20MPa; A – ion nitrided, B – pack borided, C – quenched and tempered, D – coated TiB2, E – laser borided
Fig. 3. Influence of surface treatment annular sample on moment of friction in function load of kinematics pair; A – ion nitrided, B – pack borided, C – quenched and tempered, D – coated TiB2, E – laser borided
In order to determine the effect of the test duration on the friction processes, the measurements of friction resistances were made in pre-determined load conditions (Fig. 4). The registered courses of the friction coefficient indicate a similar character of changes in the surface layer of pairs with nitrided, powder-pack borided, hardened and tempered layers. Despite the initial differences between the friction resistances after 300 seconds of operation, the level and the character of these changes is uniform, and the friction coefficient value is about 0.09. In the surface layer of other pairs, a significant increase in the friction resistance is noted; the pairs with powder-pack boronizing annular sample have the resistance coefficient of approximately 0.15, and 0.17 for the TiB2 coating. The temperature values taken in the friction areas at the end of the test fall between 80 and 90 °C and can be ranged ascending, for the annular samples used as follows: hardening and tempering, powder-pack boronizing, ion nitriding, laser boronizing and TiB2 coating .
Fig. 4. Influence of surface treatment annular sample on friction coefficient and temperature in function time of test; A – ion nitrided, B – pack borided, C – quenched and tempered, D – coated TiB2, E – laser borided
In order to determine the effect of the test duration on the friction processes, the measurements of friction resistances were made in pre-determined load conditions (Fig. 4). The registered courses of the friction coefficient indicate a similar character of changes in the surface layer of pairs with nitrided, powder-pack borided, hardened and tempered layers. Despite the initial differences between the friction resistances after 300 seconds of operation, the level and the character of these changes is uniform, and the friction coefficient value is about 0.09. In the surface layer of other pairs, a significant increase in the friction resistance is noted; the pairs with powder-pack boronizing annular sample have the resistance coefficient of approximately 0.15, and 0.17 for the TiB2 coating. The temperature values taken in the friction areas at the end of the test fall between 80 and 90 °C and can be ranged ascending, for the annular samples used as follows: hardening and tempering, powder-pack boronizing, ion nitriding, laser boronizing and TiB2 coating .
0,00 0,04 0,08 0,12 0,16 0,20
0 75 150 225 300 375 450
Time t [s]
Friction Coefficient µ
A B E D
C
20 40 60 80 100 120
0 100 200 300 400 500 Time t [s]
Temperature T [oC]
D E A B C
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Fig. 5. Influence of surface treatment annular sample on friction forces and temperature depending on load and rotation speed 100 rpm annular sample; A – ion nitrided, B – pack borided, C – quenched and tempered, D – coated TiB2, E – laser borided
Significant changes of the friction force and temperature values within the friction area occur under singular pressures p = 5, 10, 15 and 20 MPa (Fig. 5). The friction- force level values for the pairs with ion nitriding, powder-pack boronizing, hardening and tempering samples are similar and do not exceed 200 N (at 200 MPa). Its value for the laser boronizing pairs is 270 N and amounts to 341 N for the TiB2 pairs. The low- est friction force values at low-pressure conditions (at 5 MPa) were measured in pairs with nitriding and powder-pack boronizing samples. At the pressure of 10 MPa, the friction force increases by 50% in pairs with powder-pack borided surface layer in comparison to the pairs with nitrided layer. Further increase of pressure up to 15 MPa in a pair with powder-pack borided surface layer decreases the tendency toward the rise of the friction force. The hardening and tempering, laser boronizing and TiB2 coating pairs exhibit the intensity of changes within the pressure range of 5 to 10 MPa at a similar level. The temperature measurements following the conclusion of the tests have indicated the lowest heat in sliding pairs with the hardening and tempering sam- ples (below 80oC). The highest temperature values were noted for the pairs with the TiB2 coating (about 111oC) and powder-pack borided surface layer (104oC) . The change of the surface pressure affects the temperature increase within the pair’s adhe- sion area proportionally. In the nitrided and powder-pack borided surface layer pairs, the intensity of the temperature increase is smaller for the low-pressure range (5-10 MPa) than for the higher pressures (10-20 MPa). In the surface layer of all other pairs tested, however, the changes tend to be quite the opposite; higher singular pressures will decrease the intensity of the temperature increase within the friction area.
The registered courses of the friction force and temperature reveal the ability of the sliding pairs to adapt to the friction conditions in the extension of the pair’s opera- tion time. The changes occurring in the reaction of the pair to the stabilised forcing upon the start-up, and to the time flow, explain whether the system allows for a long- term and reliable operation or not. In the initial period of the pair’s operation, there is
0 100 200 300 400
A B C D E
Friction Force [N]
0 20 40 60 80 100 120
A B C D E
Temperature T [oC]
0,00 1,00
1
5 MPa 10 MPa 15 MPa 20 MPa
always an intense increase in the friction coefficient, followed by its drop and stabili- sation or increase. The stabilisation of the friction resistances indicates the adaptation of the pair composition to the existing forces and the generation of stable anti-wear and anti-seizure layers. The layers ensure the separation of the co-operating surface layer areas and a reduced rate of direct adhesion between the surface irregularities [12]. These conditions create a state of equilibrium between the processes of layer de- struction and creation within the tribochemical processes occurring in the friction pair.
The changes of the friction resistance and temperature allow assessing the probability of the kinematic sliding pair’s failure caused by the acting external forces and the emergency use of the pair [13].
These load conditions were also used for the wear measurements of the AlSn20 bearing alloy. The lowest wear was measured for pairs with nitrided, pack borided, quenched and tempered samples and did not exceed 0.01 mg of the bearing alloy’s mass, while the value dispersion was below 20% (Fig. 6). The pairs with the TiB2 coating and laser borided surface layer exhibit almost twice the wear of the AlSn20 above, amounting to 0.015 mg for the laser borided surface layer and 0.018 mg for the TiB2 coating.
The occurring differences in the wear of bearing alloy and the absence of measurable surface layer annular sample wear changes are the effect of the interaction between the co- operating surface layers, as well as of the physiochemical changes of their surfaces, induced by external forces.
These phenomena result from the elementary wear processes occurring within the contact area of the sliding pair, on the elementary surfaces of the cooperating layers. The lubrication factor is crucial for these processes, as it creates favourable or unfavourable friction conditions, depending on its transformation. These
changes contribute to the generation of boundary layers on the layers created, which are either highly resistant to ruptures or are quickly destroyed under variable operating conditions. The co-operation conditions also include the secondary phenomena of the friction and wear process. Among these are the effects of the wear products on the frictional surface layers, transmission of one element’s particles onto the other, electron emissions and the corrosion current flow [14]. The material transmission processes were observed mostly in sliding pair with pack borided surface layer and TiB2 coating (Fig. 7).
Fig. 6. Influence of surface treatment annular sample on wear of AlSn20 bearing alloy under various load conditions: A – ion nitrided, B – pack borided, C – quenched and tempered, D – coated TiB2, E – laser borided
